Rotating fluid machine

ABSTRACT

A rotating fluid machine has first energy converting means including cylinders and pistons, and second energy converting means including vanes emerging from and submerging into a rotor in the radial direction. A gaseous phase working medium having leaked out of the higher pressure first energy converting means to a reservoir is effectively utilized instead of being discarded wastefully by feeding the leaked gaseous phase working medium to the lower pressure second energy converting means via a one-way valve and a junction chamber. Moreover, when the pressure in the reservoir is lower than that in the junction chamber, the one-way valve obstructs the gaseous phase working medium from flowing back, thereby preventing the efficiency of the second energy converting means from deteriorating.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a rotating fluid machine capable offunctioning as an expander or a compressor by converting the pressureenergy of a gaseous phase working medium into the rotational energy of arotor and vice versa.

2. Description of the Related Art

Japanese Patent Laid-Open No. 2001-336491 discloses a rotating fluidmachine provided with first energy converting means composed ofcylinders and pistons and second energy converting means composed ofvanes wherein these first and second energy converting means convert thepressure energy of a gaseous phase working medium into mechanical energyto rotate a rotor and vice versa.

This rotating fluid machine includes a reservoir formed in the innercircumferential part of the rotor, in which a high pressure gaseousphase working medium having leaked out of the cylinders resides, ajunction chamber formed between the first and second energy convertingmeans for relaying the gaseous phase working medium, and an orifice forconnecting the reservoir and the junction chamber, wherein the secondenergy converting means reuses the pressure energy of the gaseous phaseworking medium having leaked to the reservoir by feeding the gaseousphase working medium to the junction chamber via the orifice.

Accordingly, it is possible to feed the gaseous phase working medium inthe reservoir to the junction chamber via the orifice when the pressurein the reservoir is higher than that in the junction chamber in theconventional rotating fluid machine. However, when the pressure in thereservoir becomes lower than that in the junction chamber, the gaseousphase working medium in the junction chamber will flow back to thereservoir via the orifice, and the pressure in the junction chamber forfeeding the gaseous phase working medium to the second energy convertingmeans will drop, inviting a problem that the energy conversionefficiency of the second energy converting means deteriorates.

SUMMARY OF THE INVENTION

An object of the present invention, attempted in view of thecircumstances stated above, is to enable a gaseous phase working mediumhaving leaked out of first or second energy converting means, whicheveris higher in pressure, to be effectively utilized by the energyconverting means lower in pressure.

In order to achieve the object stated above, according to a firstfeature of the invention, there is proposed a rotating fluid machineprovided with at least first energy converting means and second energyconverting means, capable of functioning as an expander whichintegratingly supplies mechanical energy generated by each of the firstand second energy converting means by inputting a gaseous phase workingmedium having pressure energy into the first and second energyconverting means to convert the pressure energy into mechanical energy,and capable of functioning as a compressor which integratingly suppliespressure energy of the gaseous phase working medium generated by each ofthe first and second energy converting means by inputting mechanicalenergy into the first and second energy converting means to convert themechanical energy into the pressure energy of the gaseous phase workingmedium, wherein a gaseous phase working medium having leaked out of thefirst or second energy converting means, whichever is higher inpressure, is fed to the energy converting means lower in pressure via aone-way valve.

In the configuration described above, the rotating fluid machinefunctions as an expander which integratingly supplies mechanical energygenerated by each of first and second energy converting means, andfunctions as a compressor which integratingly supplies pressure energyof the gaseous phase working medium generated by each of the first andsecond energy converting means, and the gaseous phase working mediumhaving leaked out of the energy converting means of the higher pressureis fed to the energy converting means of the lower pressure via aone-way valve. Therefore, the gaseous phase working medium having leakedout of the higher pressure energy converting means can be effectivelyutilized by the energy converting means of the lower pressure to enhancethe efficiency of energy conversion, instead of being discardedwastefully. Moreover, when the pressure of the gaseous phase workingmedium having leaked out of the higher pressure energy converting meansis lower than that of the gaseous phase working medium in the lowerpressure energy converting means, the one-way valve can obstruct thegaseous phase working medium from flowing back, thereby preventing theefficiency of the lower pressure energy converting means fromdeteriorating.

According to a second feature of the invention, in addition to the firstfeature stated above, there is proposed a rotating fluid machine whereinthe first and second energy converting means consecutively operate insuccession on a common gaseous phase working medium.

In the configuration described above, as the first and second energyconverting means consecutively operate in succession on a common gaseousphase working medium, not only can the feed and discharge routes for thegaseous phase working medium be simplified but also can the efficiencyof generating mechanical energy or pressure energy be enhanced.

According to a third feature of the invention, in addition to the firstfeature stated above, there is proposed a rotating fluid machine whereinthe first energy converting means comprises cylinders radially formed ina rotor rotatably accommodated in a rotor chamber and pistons slidingwithin these cylinders, and the second energy converting means comprisesvanes which emerge from and submerge into the rotor in the radialdirection and their outer circumferential faces are in sliding contactwith the inner circumferential face of the rotor chamber.

In the configuration described above, as the first energy convertingmeans comprises cylinders radially formed in a rotor rotatablyaccommodated in a rotor chamber and pistons sliding within thesecylinders, it is possible to minimize the efficiency drop due to leakageby enhancing the sealing performance of the high pressure gaseous phaseworking medium. Also, as the second energy converting means comprisesvanes supported by the rotor to move in the radial direction, theirouter circumferential faces being in sliding contact with the innercircumferential face of the rotor chamber, the structure of themechanism to convert pressure energy and mechanical energy is simple,making it possible to process a large quantity of gaseous phase workingmedium in spite of the compactness of the structure. This combination ofthe first energy converting means having pistons and cylinders and thesecond energy converting means having vanes makes it possible to obtaina high performance rotating fluid machine having the features of theboth.

The aforementioned and other objects, features and advantages of thepresent invention will become apparent from the following detaileddescription of the preferred embodiment thereof when taken inconjunction with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 through FIG. 14 illustrate a preferred embodiment of the presentinvention, wherein FIG. 1 is a schematic diagram outlining a waste heatrecovery device for an internal combustion engine;

FIG. 2 is a vertical sectional view of an expander taken on line 2-2 inFIG. 4;

FIG. 3 is an enlarged sectional view of a part around the axis in FIG.2;

FIG. 4 is a sectional view taken on line 4-4 in FIG. 2;

FIG. 5 is a sectional view taken on line 5-5 in FIG. 2;

FIG. 6 is a sectional view taken on line 6-6 in FIG. 2;

FIG. 7 is a sectional view taken on line 7-7 in FIG. 5;

FIG. 8 is a sectional view taken on line 8-8 in FIG. 5;

FIG. 9 is a sectional view taken on line 9-9 in FIG. 8;

FIG. 10 is a sectional view taken on line 10-10 in FIG. 3;

FIG. 11 is an exploded perspective view of a rotor; FIG. 12 is anexploded perspective view of a lubricating water distribution part ofthe rotor;

FIG. 13 is an exploded perspective view of a seal ancillary member, aspring and an end of a vane seal; and

FIG. 14 is a schematic view of sectional shapes of a rotor chamber andthe rotor.

DESCRIPTION OF THE PREFERRED EMBODIMENT

A preferred embodiment of the present invention will be described belowwith reference to the accompanying drawings.

As shown in FIG. 1, a waste heat recovery device 2 for recovering thethermal energy of exhaust gas of an internal combustion engine 1 andoutputting mechanical energy has an evaporator 3 for generating hightemperature high pressure steam by heating water with heat derived fromthe exhaust gas of the internal combustion engine 1, an expander 4 foroutputting axial torque with the expansion of the high temperature highpressure steam, a condenser 5 for cooling and liquefying reducedtemperature reduced pressure steam discharged from that expander 4, atank 6 for storing the water discharged from the condenser 5, and a lowpressure pump 7 and a high pressure pump 8 for supplying water in thetank 6 to the evaporator 3 again.

The water in the tank 6 is pressurized to 2 to 3 MPa by the low pressurepump 7 arranged on a passage P1, and preheated as it passes through aheat exchanger 102 provided in the exhaust pipe 101 of the internalcombustion engine 1. The water having passed through the heat exchanger102 and been preheated is supplied to a water jacket 105 formed in thecylinder block 103 and the cylinder head 104 of the internal combustionengine 1 via a passage P2, cools the heat generating part of theinternal combustion engine 1 as it passes therethrough, and itself isfurther heated as it deprives the heat generating part of its heat. Thewater having come out of the water jacket 105 is supplied to adistributing valve 106 via a passage P3, and is distributed therefrom toa first line communicating with a passage P4, a second linecommunicating with a passage P5, a third line communicating with apassage P6 and a fourth line communicating with passages P7.

The water distributed by the distributing valve 106 to the first lineconsisting of the passage P4 is pressurized by the high pressure pump 8to a high pressure of 10 MPa or above, and supplied to the evaporator 3,where it exchanges heat with high temperature exhaust gas to become hightemperature high pressure steam to be supplied to the high pressure partof the expander 4 (cylinders 44 . . . of the expander 4 to be describedafterwards). On the other hand, the water distributed by thedistributing valve 106 to the second line communicating with the passageP5 passes a reducing valve 107 provided on the second line to becomelower temperature lower pressure than the aforementioned hightemperature high pressure steam, and this steam is supplied to the lowpressure part of the expander 4 (the vane chambers 75 . . . of theexpander 4 to be described afterwards). In this way, as the heated waterfrom the distributing valve 106 is converted by the reducing valve 107into steam and supplied to the low pressure part of the expander 4, thewater can effectively utilize the thermal energy received in the waterjacket 105 of the internal combustion engine 1 to increase the output ofthe expander 4. The water distributed to the third line communicatingwith the passage P6 is supplied to the lubricated parts of the expander4. As the lubricated parts of the expander 4 are then lubricated byusing the high temperature water heated by the water jacket 105,overcooling of the expander 4 can be prevented to reduce the so-calledcooling loss. The reduced temperature reduced pressure steam containingthe water discharged from the expander 4 is supplied to the condenser 5provided on a passage P8, exchanges heat with a cooling air flow from acooling fan 109 driven by an electric motor 108, and the condensatewater is discharged to the tank 6. Further the water distributed to thefourth line communicating with the plurality of passages P7, is suppliedto the auxiliaries 110 such as a heater for warming the vehiclecompartment and thermoelectric elements to discharge the heat, and thewater reduced in temperature is discharged into the tank 6 via a checkvalve 111 provided on a passage P9.

The low pressure pump 7, the high pressure pump 8, the distributingvalve 106 and the electric motor 108 are controlled by an electroniccontrol unit 112 in accordance with the operating state of the internalcombustion engine 1, that of the expander 4, that of the auxiliaries 110and the temperature of water in the tank 6 and the like.

As shown in FIG. 2 and FIG. 3, the casing 11 of the expander 4 iscomposed of metallic first and second casing halves 12 and 13. The firstand second casing halves 12 and 13 consist of bodies 12 a and 13 atogether constituting a rotor chamber 14, and the round flanges 12 b and13 b integrally communicating with the outer circumference of thosebodies 12 a and 13 a. The two round flanges 12 b and 13 b are coupled toeach other by a metal gasket 15. The outer face of the first casing half12 is covered by a junction chamber outer wall 16 forming a deep bowlshape, and a round flange 16 a communicating with the outercircumference of the outer wall 16 is laid over the left side face ofthe round flange 12 b of the first casing half 12. The outer face of thesecond casing half 13 is covered by an exhaust chamber outer wall 17accommodating a magnetic coupling (not shown) for transmitting theoutput of the expander 4 outside, and a round flange 17 a integrallycommunicating with the outer circumference of the outer wall 17 is laidover the right side face of the round flange 13 b of the second casinghalf 13. The four round flanges 12 b, 13 b, 16 a and 17 a are fastenedtogether with a plurality of bolts 18 . . . arranged in thecircumferential direction. A junction chamber 19 is defined between thejunction chamber outer wall 16 and the first casing half 12, and anexhaust chamber 20 is defined between the exhaust chamber outer wall 17and the second casing half 13. The exhaust chamber outer wall 17 isprovided with an exhaust port (not shown) for guiding the reducedtemperature reduced pressure steam, which has finished its task in theexpander 4, to the condenser 5.

The bodies 12 a and 13 a of the two casing halves 12 and 13 have hollowbearing cylinders 12 c and 13 c protruding outward to the left andright, respectively, and a rotation shaft 21 having a hollow part 21 ais rotatably supported by the hollow bearing cylinders 12 c and 13 c viaa pair of bearing members 22 and 23. This causes the axis L of therotation shaft 21 to pass the intersection point of the longer axis andthe shorter axis of the substantially oval rotor chamber 14.

A seal block 25 is accommodated within a lubricating water guidingmember 24 screwed onto the right end of the second casing half 13, andfixed with a nut 26. A smaller diameter portion 21 b at the right end ofthe rotation shaft 21 is supported within the seal block 25; a pair ofseal members 27 and 27 are arranged between the seal block 25 and thesmaller diameter portion 21 b; a pair of seal members 28 and 28 arearranged between the seal block 25 and the lubricating water guidingmember 24; and further a seal member 29 is arranged between thelubricating water guiding member 24 and the second casing half 13. Afilter 30 is fitted into a concave formed in the outer circumference ofa hollow bearing cylinder 13 c of the second casing half 13, andprevented from coming off by a filter cap 31 screwed into the secondcasing half 13. A pair of seal members 32 and 33 are provided betweenthe filter cap 31 and the second casing half 13.

As is evident from FIG. 4 and FIG. 14, a circular rotor 41 is rotatablyaccommodated within the pseudo-oval rotor chamber 14. The rotor 41 isfitted onto and integrally coupled with the outer circumference of therotation shaft 21, and the axis of the rotor 41 and the axis of therotor chamber 14 coincide with each other with respect to the axis L ofthe rotation shaft 21. The shape of the rotor chamber 14, viewed in thedirection of the axis L is pseudo-oval, resembling a lozenge with itsfour corners rounded, and has a longer axis DL and a shorter axis DS.The shape of the rotor 41 viewed in the direction of the axis L iscircular, and has a diameter DR slightly shorter than the shorter axisDS of the rotor chamber 14.

The sectional shapes of both the rotor chamber 14 and the rotor 41viewed in a direction orthogonal to the axis L are like an athleticstrack. That is, the sectional shape of the rotor chamber 14 consists ofa pair of flat faces 14 a and 14 a extending in parallel with a distanced between them and arcuate faces 14 b having a central angle of 180°smoothly connecting the outer circumference of the flat faces 14 a and14 a. Similarly, the sectional shape of the rotor 41 consists of a pairof flat faces 41 a and 41 a extending in parallel with the distance dbetween them and of arcuate faces 41 b having a central angle of 180°smoothly connecting the outer circumference of the flat faces 41 a and41 a. Therefore, the flat faces 14 a and 14 a of the rotor chamber 14and the flat faces 41 a and 41 a of the rotor 41 are in contact witheach other, and a pair of crescent spaces (see FIG. 4) are formedbetween the inner circumferential face of the rotor chamber 14 and theouter circumference of the rotor 41.

Next will be described in detail the structure of the rotor 41 withreference to FIG. 3 through FIG. 6 and FIG. 11.

The rotor 41 consists of a rotor core 42 formed integrally with theouter circumference of the rotation shaft 21 and twelve rotor segments43 . . . fixed so as to cover the circumference of the rotor core 42 toconstitute a shell of the rotor 41. The twelve cylinders 44 . . . madeof ceramic (or carbon) are mounted radially to the rotor core 42 at 30°intervals, and prevented from coming off with clips 45 . . . . A smallerdiameter portion 44 a protrudes at the inner end of each of thecylinders 44, and the base end of the smaller diameter portion 44 a issealed from a sleeve 84 by a C-shaped seal 46. The tip end of thesmaller diameter portion 44 a is fitted onto the outer circumferentialface of the hollow sleeve 84, and a cylinder bore 44 b communicates withfirst and second steam passages S1 and S2 within the rotation shaft 21via third steam passages S3 . . . penetrating the smaller diameterportion 44 a and the rotation shaft 21. A ceramic piston 47 is slidablyfitted into each of the cylinders 44. When the piston 47 has movedfarthest inward in the radial direction, it is completely receded intothe cylinder bore 44 b, and when it has moved farthest outward in theradial direction, about half of its overall length protrudes out of thecylinder bore 44 b.

Each of the rotor segments 43 is a hollow wedge-shaped member having acentral angle of 30°. Two recesses 43 a and 43 b extending in an arcuateform around the axis L are formed on a face of each of the rotorsegments 43 opposite the pair of flat faces 14 a and 14 a of the rotorchamber 14. Lubricating water injection ports 43 c and 43 d open at thecenters of these recesses 43 a and 43 b. Four lubricating waterinjection ports 43 e, 43 e and 43 f, 43 f open in end faces of the rotorsegments 43, i.e. the faces opposite vanes 48 to be describedafterwards.

The rotor 41 is assembled in the following manner. The twelve rotorsegments 43 . . . are fitted onto the outer circumference of the rotorcore 42 to which the cylinders 44 . . . , the clips 45 . . . and theC-shaped seal 46 . . . are mounted in advance; and the vanes 48 . . .are fitted into twelve vane grooves 49 . . . formed between adjoiningrotor segments 43 . . . . In this process, in order form prescribedclearances between the vanes 48 . . . and the rotor segments 43 . . . ,shims of a prescribed thickness are superposed on opposite faces of thevanes 48 . . . . In this state, the rotor segments 43 . . . and thevanes 48 . . . are fastened inward in the radial directions toward therotor core 42 by using a jig and, after the rotor segments 43 . . . areprecisely positioned relative to the rotor core 42, the individual rotorsegments 43 . . . are temporarily secured to the rotor core 42 withtacking bolts 50 . . . (see FIG. 8). Then, two knock pin holes 51 and 51penetrating the rotor core 42 are co-machined with the rotor segments43, and four knock pins 52 . . . were press-fitted into those knock pinholes 51 and 51 to couple the rotor segments 43 . . . with the rotorcore 42.

As is evident from FIG. 8, FIG. 9 and FIG. 12, a through hole 53penetrating the rotor segments 43 and the rotor core 42 is formedbetween the two knock pin holes 51 and 51, and concaves 54 and 54 areformed at the opposite ends of the through hole 53. Two pipe members 55and 56 are fitted into the through hole 53 with seal members 57 through60, and an orifice forming plate 61 and a lubricating water distributingmember 62 are fitted into the respective concaves 54 and fixed with nuts63. The orifice forming plate 61 and the lubricating water distributingmember 62 are prevented from turning relative to the rotor segments 43by two knock-pins 64 and 64 penetrating knock pin holes 61 a and 61 a inthe orifice forming plate 61 and fitted into knock pin holes 62 a and 62a in the lubricating water distributing member 62. Any space between thelubricating water distributing member 62 and the nut 63 is sealed with Orings 65.

A smaller diameter portion 55 a formed at the outer end of one pipemember 55 communicates with a sixth water passage W6 within the pipemember 55 via a through hole 55 b, and the smaller diameter portion 55 acommunicates with radial distributing grooves 62 b formed in one sideface of the lubricating water distributing member 62. The distributinggrooves 62 b in the lubricating water distributing member 62 extend insix directions, and their tip ends communicate with six orifices 61 b,61 b; 61 c, 61 c; 61 d and 61 d in the orifice forming plate 61. Thestructures of the orifice forming plate 61, the lubricating waterdistributing member 62 and the nut 63 provided at the outer end of theother pipe member 56 are respectively the same as those of the orificeforming plate 61, the lubricating water distributing member 62 and thenuts 63 described above.

The two orifices 61 b and 61 b of the orifice forming plate 61communicate on the downstream side with the lubricating water injectionports 43 e and 43 e opening to be opposite the vanes 48 via seventhwater passages W7 and W7 formed within the rotor segments 43, two otherorifices 61 c and 61 c communicate on the downstream side with theaforementioned two lubricating water injection ports 43 f and 43 fopening to be opposite the vanes 48 via eighth water passages W8 and W8formed within the rotor segments 43, and the two still other orifices 61d and 61 d communicate on the downstream side with the aforementionedtwo lubricating water injection ports 43 c and 43 d opening to beopposite the rotor chamber 14 via ninth water passages W9 and W9 formedwithin the rotor segments 43.

As is evident when FIG. 5 is referenced together, an annular groove 67partitioned by the pair of O rings 66 and 66 is formed in the outercircumference of the cylinders 44, and the sixth water passage W6 formedin one pipe member 55 communicates with the annular groove 67 via fourthrough holes 55 c . . . penetrating the pipe member 55 and a 10th waterpassage W10 formed within the rotor core 42. The annular groove 67communicates with the sliding faces of the cylinder bore 44 b and thepiston 47 via the orifices 44 c. The position of the orifice 44 c ineach cylinder 44 is set where it does not deviate from the sliding faceof the piston 47 when the piston 47 moves between its top dead centerand bottom dead center.

As is evident from FIG. 3 and FIG. 9, a first water passage W1 formed inthe lubricating water guiding member 24 communicates with the smallerdiameter portion 55 a of the one pipe member 55, via second waterpassage W2 formed in the seal block 25, third water passages W3 . . .formed in the smaller diameter portion 21 b of the rotation shaft 21, anannular groove 68 a formed in the outer circumference of a water passageforming member 68 fitted into the core of the rotation shaft 21, afourth water passage W4 formed in the rotation shaft 21, a pipe member69 provided astride over the rotor core 42 and the rotor segments 43,and fifth water passages W5 and W5 formed so as to detour a knock pin 52inside the rotor segments 43 in the radial direction.

As shown in FIG. 5, FIG. 7, FIG. 9 and FIG. 11, twelve vane grooves 49 .. . extending in radial directions are formed between adjoining rotorsegments 43 . . . of the rotor 41, and the plate-shaped vanes 48 . . .are fitted into these vane grooves 49 . . . to be slidable. Each of thevanes 48 is formed in a substantially U shape provided with parallelfaces 48 a and 48 a positioned along the parallel faces 14 a and 14 a ofthe rotor chamber 14, an arcuate face 48 b along an arcuate face 14 b ofthe rotor chamber 14, and a notch 48 c positioned between the twoparallel faces 48 a and 48 a, and rollers 71 and 71 each having a rollerbearing structure are rotatably supported by a pair of spindles 48 d and48 d protruding from the two parallel faces 48 a and 48 a.

Slit-shaped seal holding grooves 48 f are formed from the arcuate face48 b of each vane 48 to the pair of the parallel faces 48 a and 48 a.Each of these seal holding grooves 48 f holds a synthetic resin-madevane seal 72 formed in a U shape, and the tip of this vane seal 72slightly protrudes from the outer circumferential face of the vane 48 tocome into sliding contact with the inner circumferential face of therotor chamber 14. Engaging holes 48 g and 48 g having a circular sectioncommunicating with the inner ends of the seal holding grooves 48 f inthe radial direction are formed in the direction of the axis L in thepair of parallel faces 48 a and 48 a of the vane, and cylindrical sealancillary members 76 and 76 are fitted into these engaging holes 48 gand 48 g with no gaps therebetween. As is evident from FIG. 13, slits 76a and 76 a opening outward in the radial directions and outward in theaxial direction are formed in the seal ancillary members 76 and 76, andthe inner ends of the vane seals 72 in the radial direction are fittedinto these slits 76 a and 76 a with no gaps therebetween. The sealancillary members 76 and 76 are pressed outward in the direction of theaxis L (the direction of protruding out of the engaging holes 48 g and48 g) by springs 77 and 77 arranged on the bottoms of the engaging holes48 g and 48 g.

Formed on the opposite sides of each vane 48 are two recesses 48 e and48 e, and these recesses 48 e and 48 e are opposite the two lubricatingwater injection ports 43 e and 43 e inside in the radial direction,opening in the end faces of the rotor segments 43. Formed within thevane 48 is a trap chamber 48 h extending in the radially inward andoutward directions; the inside of the trap chamber 48 h in the radialdirection communicates, via suction ports 48 i and 48 i opening in thetwo sides of the vane 48, with a reservoir 78 formed between the rotorcore 42 and the rotor segments 43 . . . , and the outside of the trapchamber 48 h in the radial direction communicates, via an exhaust port48 j opening in the forward side face of the vane 48 in the rotationaldirection R, with the vane chamber 75. A piston bearing member 73protruding at the center of the notch 48 c in the vane 48 in theradially inward direction comes into contact with the outer end of thepiston 47 in the radial direction.

As is evident from FIG. 2, the reservoir 78 formed between the rotorcore 42 and the rotor segments 43 . . . and the junction chamber 19 aremade communicable with each other by a communicating hole 12 dpenetrating the first casing 12, and a one-way valve 79 permitting theshift of steam from the reservoir 78 to the junction chamber 19 andrestricting the shift of steam from the junction chamber 19 to thereservoir 78 is arranged in this communicating hole 12 d.

As is evident from FIG. 4, pseudo-oval annular grooves 74 and 74resembling diamonds with the four corners rounded are provided in theflat faces 14 a and 14 a of the rotor chamber 14 partitioned by thefirst and second casing halves 12 and 13, and the pair of rollers 71 and71 of each vane 48 engage to be able to roll in the two annular grooves74 and 74. The distance between these annular grooves 74 and the arcuatefaces 14 b of the rotor chamber 14 is constant over the entirecircumference. Therefore, when the rotor 41 turns, the vanes 48, whoserollers 71 and 71 are guided by the annular grooves 74 and 74,reciprocate in the radial direction within the vane grooves 49 and, in astate in which the vane seal 72 mounted to the arcuate face 48 b of eachvane 48 is compressed to a certain amount, slide along the arcuate faces14 b of the rotor chamber 14. This results in preventing the rotorchamber 14 and the vanes 48 . . . from coming into direct solid contact,and makes it possible to reliably seal the vane chambers 75 . . .partitioned by the adjoining vanes 48 . . . while preventing an increasein frictional resistance and the occurrence of wear.

As is evident from FIG. 2, FIG. 3 and FIG. 10, an aperture 16 b isformed at the center of the junction chamber outer wall 16, and a bossportion 81 a of a fixed shaft supporting member 81 arranged on the axisL is fixed to the inner face of the aperture 16 b with a plurality ofbolts 82 . . . and further fixed to the first casing half 12 with a nut83. The ceramic cylindrical sleeve 84 is fixed to the hollow part 21 aof the rotation shaft 21, and the outer circumferential face of a fixedshaft 85 integral with the fixed shaft supporting member 81 is fittedonto the inner circumferential face of this sleeve 84 to be rotatable ina relative manner. Any space between the left end of the fixed shaft 85and the first casing half 12 is sealed by a seal member 86, while anyspace between the right end of the fixed shaft 85 and the rotation shaft21 is sealed by a seal member 87.

A steam feed pipe 88 is fitted into the fixed shaft supporting member 81arranged on the axis L and fixed with a nut 89, and the right end ofthis steam feed pipe 88 is pressed into the core of the fixed shaft 85.The first steam passages S1 communicating with the steam feed pipe 88 isformed in the axial direction at the core of the fixed shaft 85, and thepair of second steam passages S2 and S2 penetrate the fixed shaft 85 inthe radial direction with a phase difference of 180°. As stated above,the twelve third steam passages S3 . . . penetrate the sleeve 84 and thesmaller diameter portions 44 a . . . of the 12 cylinders 44 . . . heldat 30° intervals by the rotor 41 fixed to the rotation shaft 21, and theinner ends of these third steam passages S3 . . . in the radialdirection are communicatably opposite the outer ends of the second steampassages S2 and S2 in the radial direction.

A pair of notches 85 a and 85 a are formed in the outer circumferentialface of the fixed shaft 85 with a phase difference of 180°, and thesenotches 85 a and 85 a are communicatable with the third steam passagesS3 . . . . The notches 85 a and 85 a and the junction chamber 19communicate with each other via a pair of fourth steam passages S4 andS4 formed in the fixed shaft 85 in the axial direction, an annular fifthsteam passage S5 formed in the fixed shaft supporting member 81 in theaxial direction, and through holes 81 b . . . opening in the outercircumference of the boss portion 81 a of the fixed shaft supportingmember 81.

As shown in FIG. 2 and FIG. 4, a plurality of air intake ports 90 . . .aligned in radial directions are formed in the first casing half 12 andthe second casing half 13 in 15° forward positions in the rotationaldirection R of the rotor 41 with reference to the direction of theshorter axis of the rotor chamber 14. These air intake ports 90 . . .make the internal space of the rotor chamber 14 communicate with thejunction chamber 19. In the second casing half 13 are formed a pluralityof exhaust ports 91 . . . in 15° to 75° backward positions in therotational direction R of the rotor 41 with reference to the directionof the shorter axis of the rotor chamber 14. These exhaust ports 91 . .. make the internal space of the rotor chamber 14 communicate with theexhaust chamber 20. In order that the vane seals 72 . . . of the vanes48 . . . may not be damaged by the edges of the exhaust ports 91 . . . ,those exhaust ports 91 . . . open into shallow concaves 13 d and 13 dformed inside the second casing half 13.

The second steam passages S2 and S2, the third steam passages S3 . . .the notches 85 a and 85 a of the fixed shaft 85 and the third steampassages S3 . . . constitute a rotary valve V which is periodically madecommunicable by the relative rotations of the fixed shaft 85 and therotation shaft 21 (see FIG. 10).

As is evident from FIG. 2, an 11th water passage W11 formed in the firstand second casing halves 12 and 13 communicates with the outercircumferential face of the annular filter 30 via a 14th water passageW14 comprising a pipe, and the inner circumferential face of the filter30 communicates with a 16th water passage W16 formed in the secondcasing half 13 via a 15th water passage W15 formed in the second casinghalf 13. Water fed to the 16th water passage W16 lubricates the slidingfaces of the fixed shaft 85 and the sleeve 84. Water fed from the innercircumferential face of the filter 30 to the outer circumference of thebearing member 23 via a 17th water passage W17 lubricates the outercircumferential face of the rotation shaft 21 through an orificepenetrating the bearing member 23. On the other hand, water fed from the11th water passage W11 to the outer circumference of the bearing member22 via an 18th water passage W18 comprising a pipe, after lubricatingthe outer circumferential face of the rotation shaft 21 through anorifice penetrating the bearing member 22, lubricates the sliding facesof the fixed shaft 85 and the sleeve 84.

Next will be described the operations of this embodiment of theinvention configured as described above.

First will be described the operation of the expander 4. With referenceto FIG. 3, high temperature high pressure steam from the evaporator 3 isfed to the steam feed pipe 88, the first steam passage S1 penetratingthe core of the fixed shaft 85, and the pair of second steam passages S2and S2 penetrating the fixed shaft 85 in the radial direction. Referringto FIG. 10, when the sleeve 84 rotating integrally with the rotor 41 andthe rotation shaft 21 in the direction of arrow R reaches apredetermined phase relative to the fixed shaft 85, the pair of thirdsteam passages S3 and S3 in a forward position in the rotationaldirection R of the rotor 41 from the position of the shorter diameter ofthe rotor chamber 14 come to communicate with the pair of second steampassages S2 and S2, and the high temperature high pressure steam of thesecond steam passages S2 and S2 is fed into the insides of the pair ofcylinders 44 and 44 via the third steam passages S3 and S3, and pressesthe pistons 47 and 47 outward in the radial directions. Referring toFIG. 4, when the vanes 48 and 48 pressed by these pistons 47 and 47 moveoutward in the radial direction, engagement between the pair of rollers71 provided on the vanes 48 and the annular grooves 74 and 74 convertsthe forward movements of the pistons 47 and 47 into a rotationalmovement of the rotor 41.

Even after the communication between the second steam passages S2 and S2and the third steam passages S3 and S3 is cut off due to the rotation ofthe rotor 41, the high temperature high pressure steam in the cylinders44 and 44 continues to expand and causes the pistons 47 and 47 to movefarther forward, thereby continuing the rotation of the rotor 41. Whenthe vanes 48 and 48 reaches the position of the longer diameter of therotor chamber 14, the third steam passages S3 and S3 communicating withthe respectively corresponding cylinders 44 and 44 come to communicatewith the notches 85 a and 85 a of the fixed shaft 85, and the pistons 47and 47 pressed by the vanes 48 and 48 whose rollers 71 and 71 are guidedby the annular grooves 74 and 74 moves in the radially inward directionand causes steam in the cylinders 44 and 44 to pass through the thirdsteam passages S3 and S3, the notches 85 a and 85 a, the fourth steampassages S4 and S4, the fifth steam passage S5 and the through holes 81b . . . to become first reduced temperature reduced pressure steam,which is fed to the junction chamber 19. The first reduced temperaturereduced pressure steam results from the temperature and pressurereduction of the high temperature high pressure steam fed from the steamfeed pipe 88 and having finished the work to drive the pistons 47 and47. The thermal energy and pressure energy of the first reducedtemperature reduced pressure steam are still sufficient to drive thevanes 48 . . . , though weaker than those of the high temperature highpressure steam.

The first reduced temperature reduced pressure steam in the junctionchamber 19 is fed from the air intake ports 90 . . . of the first andsecond casing halves 12 and 13 to the vane chambers 75 . . . within therotor chamber 14, where it further expands to press the vanes 48 . . .thereby to turn the rotor 41. The second reduced temperature reducedpressure steam having finished its task to have lowered temperature andpressure is discharged from the exhaust ports 91 . . . of the secondcasing half 13 into the exhaust chamber 20 and fed from there to thecondenser 5.

As the twelve pistons 47 . . . are successively actuated by theexpansion of high temperature high pressure steam in this way to turnthe rotor 41 via the rollers 71 and 71 and the annular grooves 74 and74, and the rotor 41 is turned via the vanes 48 . . . by the expansionof the first reduced temperature reduced pressure steam resulting from adecrease in temperature and pressure of the high temperature highpressure steam, it is possible to integrate mechanical energy generatedby the pistons 47 . . . and mechanical energy generated by the vanes 48. . . to obtain an output from the rotation shaft 21 and, moreover, thepressure energy of the high temperature high pressure steam can becompletely converted into mechanical energy.

Further, as first energy converting means is composed of the cylinders44 . . . radially formed in the rotor 41 rotatably accommodated withinthe rotor chamber 14 and the pistons 47 . . . sliding inside thesecylinders 44 . . . , it is made possible to minimize the efficiency dropdue to leaks by enhancing the sealing performance of the hightemperature high pressure gaseous phase working medium. Also, as secondenergy converting means is composed of the vanes 48 . . . supported bythe rotor 41 to be movable in the radial directions and being in slidingcontact with the inner circumferential face of the rotor chamber 14, thestructure of the mechanism to convert pressure energy into mechanicalenergy is simple, thereby processing a large quantity of gaseous phaseworking medium in spite of the compact structure. Moreover, thecombination of the first energy converting means having the cylinders 44. . . and the pistons 47 . . . with the second energy converting meanshaving the vanes 48 . . . results in a high performance rotating fluidmachine having the features of the both.

Next will be described the aforementioned lubrication of the vanes 48 .. . and the pistons 47 . . . of the expander 4 with water.

For lubricating each part of the expander 4, there is used hightemperature water distributed by the distributing valve 106 to thepassage P6 after being heated by the water jacket 105.

Referring to FIG. 3 and FIG. 8, water supplied to the first waterpassage W1 of the lubricating water guiding member 24 flows into thesmaller diameter portion 55 a of the one pipe member 55 via the secondwater passage W2 . . . of the seal block 25, the third water passages W3. . . of the rotation shaft 21, the annular groove 68 a of the waterpassage forming member 68, the fourth water passage W4 of the rotationshaft 21, the pipe member 69, and the fifth water passages W5 and W5formed in the rotor segments 43. Water fed into the smaller diameterportion 55 a flows into a smaller diameter portion 56 a of the otherpipe member 56 via the through hole 55 b in the one pipe member 55, thesixth water passage W6 formed in the two pipe members 55 and 56, and athrough hole 56 b formed in the other pipe member 56.

Water from the respective smaller diameter portions 55 a and 56 a of thepipe members 55 and 56 passes through the six orifices 61 b, 61 b; 61 c,61 c; 61 d and 61 d in the orifice forming plate 61 via the distributinggrooves 62 b of the lubricating water distributing member 62, one partof the water is injected out of the four lubricating water injectionports 43 e, 43 e; 43 f, 43 f opening in the end faces of the rotorsegments 43, and the other part is injected out of the lubricating waterinjection ports 43 c and 43 d in the arcuate recesses 43 a and 43 bformed on side faces of the rotor segments 43.

Then, the water injected out of the lubricating water injection ports 43e, 43 e; 43 f, 43 f of the respective end faces of the rotor segments 43into the vane grooves 49 constitutes a static pressure bearing betweenthe vane grooves 49 and the vanes 48 slidably fitted into the vanegrooves 49 to support the vanes 48 in a floating state, and preventsstatic contact between the end faces of the rotor segments 43 and thevanes 48 thereby to prevent seizure and wear from occurring. Bysupplying water to lubricate the sliding faces of the vanes 48 via waterpassages arranged radially within the rotor 41, not only can the waterbe pressurized by a centrifugal force but also the impact of thermalexpansion is reduced by stabilizing the temperature around the rotor 41and steam leakage is minimized by maintaining the preset clearances.

Further, as water is held by the two recesses 48 e and 48 e formed ineach of the opposite faces of the vanes 48, these recesses 48 e and 48 eserve as pressure reservoirs to reduce a pressure drop due to waterleakage. As a result, the vanes 48 positioned between the end faces ofthe pair of rotor segments 43 and 43 are kept in a floating state by thewater, and the frictional resistance can be effectively reduced.Although the reciprocation of the vanes 48 varies the positions of thevanes 48 relative to the rotor 41 in the radial direction, thereciprocating vanes 48 are held in a floating state all the time toenable an effective reduction in frictional resistance, because therecesses 48 e and 48 e are arranged not on the rotor segment 43 side buton the vane 48 side and disposed near the rollers 71 and 71, where theload on the vanes 48 is the greatest.

When the individual vanes 48 turn together with the rotor 41, the vaneseal 72 fitted into the seal holding grooves 48 f are pressed by acentrifugal force outward in the radial direction, and this causes thevane seal 72 to be pressed against the inner circumferential face of therotor chamber 14 in the portions of the vanes 48 matching the arcuateface 48 b, thereby exhibiting a sealing performance. Although nopressing force of the vane seal 72 due to a centrifugal force can beexpected in the portions of the vanes 48 . . . matching the parallelfaces 48 a and 48 a, the pressure introduced from the vane chamber 75into the bottom part of the seal holding grooves 48 f of the vanes 48presses the vane seal 72 in the direction of being thrust out of theseal holding grooves 48 f, so that the whole area of the outercircumferential face of the vane seal 72 is pressed against the innercircumferential face of the rotor chamber 14, thereby exhibiting asealing performance.

In this process, if the pressure escapes from the opposite ends of theseal holding grooves 48 f, the-pressing force of the vane seal 72 willgenerally disappear, but in this embodiment, the ends of the vane seal72 are fitted into the slits 76 a and 76 a of the seal ancillary members76 and 76 which are fitted into the engaging holes 48 g and 48 gcommunicating with the opposite ends of the seal holding grooves 48 f,the slits 76 a and 76 a of the seal ancillary members 76 and 76 openoutward in the radial direction and close inward in the radialdirection, and the outer end faces of the seal ancillary members 76 and76 in the direction of the axis L where the slits 76 a and 76 a open arepressed by the urging forces of the springs 77 and 77 toward the innercircumferential face of the rotor chamber 14, resulting in that the endsof the vane seal 72 are stuck tightly to the slits 76 a and 76 a of theseal ancillary members 76 and 76, whereby the sealing performance of thevane seal 72 can be ensured by preventing pressure from escaping out ofthe opposite ends of the seal holding grooves 48 f.

Especially when the expander 4 is cold and the pressure in the bottompart of the seal holding grooves 48 f does not sufficiently rise, thesealing performance can be secured by causing the urging forces of thesprings 77 and 77 to press the seal ancillary members 76 and 76 and theends of the vane seal 72 against the inner circumferential face of therotor chamber 14.

Further with reference to FIG. 5, water fed from the sixth water passageW6 within the pipe members 55 to the sliding faces of the cylinders 44and the pistons 47 via a 10th water passage W10 within the rotorsegments 43 and the annular grooves 67 in the outer circumferences ofthe cylinders 44 performs a sealing performance by virtue of theviscosity of the water film formed on those sliding faces, andeffectively prevents the high temperature high pressure steam suppliedto the cylinders 44 from leaking through space between their slidingfaces and the pistons 47. As water then supplied through the inside ofthe expander 4, which is in a high temperature state, to the slidingfaces of the cylinders 44 and the pistons 47 is heated, the hightemperature high pressure steam supplied to the cylinders 44 is cooledby that water, making it possible to minimize the output drop of theexpander 4.

Further, the first water passage W1 and the 11th water passage W11 areindependent of each other, and supply water under the pressure requiredby each part to be lubricated. More specifically, since the watersupplied from the first water passage W1 mainly supports the vanes 48 .. . and the rotor 41 in a floating state by a static pressure bearing,it needs a pressure high enough to be able to antagonize loadvariations. Unlike that, as the water supplied from the 11th waterpassage W11 mainly lubricates the surroundings of the fixed shaft 85 andseals the high temperature high pressure steam leaking from the thirdsteam passages S3 and S3 to the outer circumference of the fixed shaft85 to thereby reduce the effects of the thermal expansion of the fixedshaft 85, the rotation shaft 21, the rotor 41 and the like, the watermerely needs a pressure at least higher than that in the junctionchamber 19.

As described above, there are thus provided two water feed linesincluding the first water passage W1 for supplying high pressure waterand the 11th water passage W11 for supplying water of a lower pressurethan that, which eliminate the problems in the case where only one waterfeed line is provided. Thus, it is possible to prevent the flow rate ofwater to the junction chamber 19 from being increased by the supply ofwater under excessive pressure to the surroundings of the fixed shaft 85and the steam temperature from being lowered by excessive cooling of thefixed shaft 85, the rotation shaft 21, the rotor 41 and so forth, andthus to raise the output of the expander 4 while reducing the quantityof water supply.

Moreover, since water, which is the same substance as steam, is used asthe sealing medium, no problem will arise even if the water gets mixedwith steam. If the sliding faces of the cylinders 44 and the pistons 47were sealed with oil, the oil would inevitably get mixed with water orsteam, so that a special filtering device to separate the oil isrequired. Also, as part of the water for lubricating the sliding facesof the vanes 48 and the vane grooves 49 is bypassed to be diverted forsealing the sliding faces of the cylinders 44 and the pistons 47, thereis no need to separately provide water passages to guide that water tothe sliding faces, so that the structure can be simplified.

The liquid phase working medium supplied to the sliding faces of thevanes 48 and the vane grooves 49 to constitutes the static pressurebearings resides in the reservoir 78 formed between the rotor core 42and the rotor segments 43 . . . after having completed its role. Sincethe annular grooves 74 and 74 provided in the vanes 48 to guide therollers 71 and 71 communicate with this reservoir 78, the liquid phaseworking medium having flowed into the annular grooves 74 and 74generates a great resistance when the rollers 71 and 71 shift, whichmight invite a drop in the output of the expander 4.

However, in this embodiment, the liquid phase working medium in thereservoir 78 can be discharged to the exhaust ports 91 . . . via thevane chamber 75 by the function of the trap chambers 48 h provided onthe vanes 48. That is, as shown on the right side of FIG. 5, when thevanes 48 have retreated deepest inside the vane grooves 49, the suctionports 48 i and 48 i communicating with the inner ends of the trapchambers 48 h in the radial direction communicate with the reservoir 78,and cause the liquid phase working medium in the reservoir 78 to betrapped by the trap chambers 48 h. When the rotor 41 turns in thedirection of arrow R, the vanes 48 protrude outside the vane grooves 49in the radial direction as shown in the lower side of FIG. 5, and theexhaust ports 48 j communicating with the outer ends of the trapchambers 48 h in the radial direction communicate with the vane chamber75 in the exhaust stroke, and cause the liquid phase working mediumtrapped by the trap chambers 48 h to be discharged into the vane chamber75.

Along with the rotation of the rotor 41 in the direction of arrow R inthis way, the trap chamber 48 h provided in each of the vanes 48 causesthe liquid phase working medium in the reservoir 78 to be dischargedinto the vane chamber 75 to prevent the rotation of the rotor 41 frombeing braked by the resistance of the liquid phase working mediumresided in the reservoir 78. Moreover, when the suction ports 48 i and48 i communicate with the reservoir 78, the exhaust port 48 j does notcommunicate with the vane chamber 75, and when the exhaust port 48 jcommunicates with the vane chamber 75, the suction ports 48 i and 48 ido not communicate with the reservoir 78, in other words, the suctionports 48 i and 48 i and the exhaust port 48 j never communicate with thereservoir 78 and the vane chamber 75 at the same time. Therefore, thehigh temperature high pressure steam with pressure energy having leakedout of the sliding faces of the cylinders 44 and the pistons 47 andtrapped by the reservoir 78 is never wastefully discarded into the vanechamber 75 via the trap chamber 48 h.

Further, since the high temperature high pressure steam with pressureenergy having leaked out of the sliding faces of the cylinders 44 andthe pistons 47 and trapped by the reservoir 78 is supplied to thejunction chamber 19 via the communicating hole 12 d of the first casing12 and the one-way valve 79 (see FIG. 2), the high temperature highpressure steam can be fed from the air intake ports 90 . . . to the vanechambers 75 . . . for effective reuse. If the pressure in the reservoir78 drops below that in the junction chamber 19 for some reason, theone-way valve 79 closes to prevent reduced temperature reduced pressuresteam in the junction chamber 19 from flowing back to the reservoir 78.Therefore, it is possible to obstruct pressure from escaping out of thejunction chamber 19, thereby preventing the efficiency of the expander 4from deteriorating.

Next will be described the operation of the cooling system of theinternal combustion engine 1 including the waste heat recovery device 2mainly with reference to FIG. 1 and FIG. 2.

Water drawn up from the tank 6 by the low pressure pump 7 is fed to theheat exchanger 102 provided in the exhaust pipe 101 via the passage P1and, after being preheated there, is fed to the water jacket 105 of theinternal combustion engine 1 via the passage P2. Water flowing withinthe water jacket 105 cools the cylinder blocks 103 and the cylinderheads 104, which are the heat generating parts of the internalcombustion engine 1, and fed to the distributing valve 106 in a state ofbeing raised in temperature. As water preheated by the heat exchanger102 of the exhaust pipe 101 is fed to the water jacket 105, when theinternal combustion engine 1 is at low temperature, its warming-up canbe accelerated, and the performance of the evaporator 3 can be enhancedby preventing the internal combustion engine 1 from being excessivelycooled to raise its exhaust gas temperature.

Part of the high temperature water distributed by the distributing valve106 is pressurized by the high pressure pump 8 provided on the passageP4 and fed to the evaporator 3, where it exchanges heat with exhaust gasto become high temperature high pressure steam. The high temperaturehigh pressure steam generated by the evaporator 3 is fed to the steamfeed pipe 88 of the expander 4 and, after passing by the cylinders 44 .. . and the vane chambers 75 . . . to drive the rotation shaft 21, isdischarged into the condenser 5.

Another part of the high temperature water distributed by thedistributing valve 106 is reduced in pressure by the reducing valve 107provided on the passage P5 to become steam, which is fed to the junctionchamber 19 of the expander 4. The steam fed to the junction chamber 19joins the first reduced temperature reduced pressure steam fed from thesteam feed pipe 88 and having passed by the cylinders 44 . . . and,after driving the rotation shaft 21, is discharged to the condenser 5.As part of the high temperature water from the distributing valve 106 isvaporized by the reducing valve 107 and fed to the expander 4 in thisway, the thermal energy that the water has received from the waterjacket 105 of the internal combustion engine 1 can be effectivelyutilized to boost the output of the expander 4. Also, still another partof the high temperature water distributed by the distributing valve 106is fed to the first water passage W1 of the expander 4 via the passageP6 and lubricates various parts to be lubricated. As the parts to belubricated of the expander 4 are thus lubricated by using hightemperature water, it is possible to prevent the expander 4 from beingexcessively cooled thereby reducing so-called cooling loss. Further, thewater having entered after lubrication into the vane chambers 75 . . .in the expansion stroke is heated and vaporized by being mixed withsteam in the vane chambers 75 and its expansion effect boosts the outputof the expander 4. Then, the second reduced temperature reduced pressuresteam discharged from the expander 4 to the passage P8 is fed to thecondenser 5, and cooled there by the cooling fan 109 to become water,which is returned to the tank 6. Still another part of the hightemperature water distributed by the distributing valve 106, after beingcooled by heat exchanging with the auxiliaries 110 provided on thepassage P7, is returned to the tank 6 via the check valve 111.

As described above, a water circulating route by which, after feedingthe water jacket 105 with water drawn up from the tank 6 by the lowpressure pump 7 to cool the heat generating part of the internalcombustion engine 1, the water is returned to the tank 6 after being fedto the auxiliaries 110 to be cooled, and a water circulating route ofthe waste heat recovery device 2 by which part of water having come outof the water jacket 105 is distributed as the working medium and thewater is returned to the tank 6 via the high pressure pump 8, theevaporator 3, the expander 4 and the condenser 5 are combined, and thewater circulating route of the cooling system for the internalcombustion engine 1 passing through the water jacket 105 and theauxiliaries 110 is provided with a low pressure and a high flow rate,while the water circulating route of the waste heat recovery device 2 isprovided with a high pressure and a low flow rate. Therefore, it ispossible to supply water at respectively suitable flow rates andpressures to the cooling system of the internal combustion engine 1 andthe waste heat recovery device 2, thereby sufficiently cooling the heatgenerating part of the internal combustion engine 1 to dispense with aradiator while maintaining the performance of the waste heat recoverydevice 2. Moreover, since water fed from the low pressure pump 7 to thewater jacket 105 is preheated by the heat exchanger 102 disposed in theexhaust pipe 101, the waste heat of the internal combustion engine 1 canbe utilized even more effectively.

Also, as the heat exchanger 102 to which low temperature water is fedfrom the low pressure pump 7 is arranged downstream from the exhaustpipe 101 where the temperature of exhaust gas is lower than in theposition of the evaporator 3, the surplus waste heat held by the exhaustgas can be recovered thoroughly and efficiently. Furthermore, as thewater preheated by the heat exchanger 102 is fed to the water jacket105, excessive cooling of the internal combustion engine 1 can beprevented and, at the same time, it is possible to further raise thecombustion heat, namely the temperature of the exhaust gas to increaseits thermal energy and to enhance the efficiency of waste heat recovery.

Although the preferred embodiment of the present invention have beendescribed in detail so far, the invention can be modified in design invarious ways without deviating from the subject matter.

For instance, although the expander 4 is described as an example ofrotating fluid machines in this embodiment, the invention is alsoapplicable to a compressor.

Although this embodiment uses steam and water as the gaseous phaseworking medium and the liquid phase working medium, respectively, anyother appropriate working medium may be used as well.

1. A rotating fluid machine provided with at least first energyconverting means and second energy converting means, capable offunctioning as an expander which integratingly supplies mechanicalenergy generated by each of the first and second energy converting meansby inputting a gaseous phase working medium having pressure energy intothe first and second energy converting means to convert the pressureenergy into mechanical energy, and capable of functioning as acompressor which integratingly supplies pressure energy of the gaseousphase working medium generated by each of the first and second energyconverting means by inputting mechanical energy into the first andsecond energy converting means to convert the mechanical energy into thepressure energy of the gaseous phase working medium, wherein a gaseousphase working medium having leaked out of the first or second energyconverting means, whichever is higher in pressure, is fed to the energyconverting means lower in pressure via a one-way valve.
 2. The rotatingfluid machine according to claim 1, wherein the first and second energyconverting means consecutively operate in succession on a common gaseousphase working medium.
 3. The rotating fluid machine according to claim1, wherein the first energy converting means comprises cylindersradially formed in a rotor rotatably accommodated in a rotor chamber andpistons sliding within these cylinders, and the second energy convertingmeans comprises vanes which emerge from and submerge into the rotor inthe radial direction and their outer circumferential faces are insliding contact with the inner circumferential face of the rotorchamber.